Positive-displacement, reciprocating compressors generally operate by increasing the pressure value of a gaseous fluid through the mechanical energy drawn from an electrical motor or a combustion engine.
Compressors based on the classical crank mechanism (see FIG. 1) for converting a rotational motion of a motor into a rectilinear reciprocating motion, have various drawbacks, the most important of which are:                The amount of frictional force, shortened as “Fia”, which adds to the force due to the action of the gases on the seals (piston rings or packing rings), and which acts between the piston side wall and the cylinder wall during the sliding of the piston, because of the reaction to the thrust exerted by the obliquity of the piston rod (connecting rod);        the overturning action exerted by the piston rod on the piston, for which reason the latter usually has a sufficient length to limit this action and to reduce the risk of seizure, thus causing, however, a dimensional and weight increase with a concomitant increase of inertial forces.        The law of motion of the piston is not perfectly sinusoidal but contains harmonics of higher order, and this causes the well known balance difficulties. These harmonics, including the lowest order one, cannot simply be balanced by counterweights; instead, they require the utilisation of counter-rotating shafts. Actually, a principle of the background art that would brilliantly solve the inherent problems of the conventional crank mechanism, is shown in FIGS. 2 and 3 and in FIG. 4 and 5.        
In this crank mechanism, by imposing a rotation on the shaft with trace O (planet carrier), the element ΩB (pinion) will move in such a way that point B will displace itself along the cylinder axis in a rectilinear manner. Several known techniques have put into practice the just described mechanism (called from now on “non-conventional”, though already known, only to distinguish it from the classical crank mechanism), but nonetheless, they have not been successful since they offer technical solutions that have some inconsistencies and prevent a correct operation, while in other cases they result in a great structural complexity which discourages their use.
As matter of fact, this technology has not been converted into an effective industrial application, notwithstanding the fact that some solutions appear to be valid; this is due to the complex structure, and to space and reliability problems, which render this system less competitive than the classical crank mechanism in the configurations proposed until now.
Summing up, this “non-conventional”, or “non-classical” crank mechanism, which is schematically shown in FIGS. 2, 3, 4, 5 and which has been adopted by the present invention, but which has been further improved in a way to be described later, has the following features.
We start with the classical crank mechanism (FIG. 1) and split the connecting rod (piston rod) OB in two identical parts, thereby obtaining two cranks OΩ and ΩB (FIG. 2). By imposing on the crank OΩ an anticlockwise rotation, and on the crank ΩB an identical but opposite rotation
“-α”, point B necessarily moves rectilinearly along the cylinder axis.
Thus, the angle formed between the connecting rod and the cylinder axis is constantly equal to zero, and consequently, the component of the forces “N”, normal to this axis, which are due to the connecting rod obliquity, reduces to zero. On the other hand, since no relative rotation exists between the connecting rod and the piston, there is no need, anymore, to provide a hinged connection at point C as in the classical crank mechanism; in other words, the gudgeon pin can be eliminated altogether and the connecting rod may be integrally formed with the piston. From the point of view of their practical realisation, the motions of the crank OΩ and of the auxiliary crank ΩB may be obtained using a pair of gearwheels, one of which has an inner toothing, centre O, is fixed with respect to a frame and has a pitch diameter 2r, while the second gearwheel has an external toothing with pitch diameter r, it meshes with the first gearwheel, and rotates around the axis passing through Ω which is integral with the crank (FIG. 3). Two possible practical realisations of this “non-conventional” crank mechanism are respectively shown in FIG. 4 and FIG. 5. This is actually a particular planetary gear train (FIG. 6) in which the central gear (the sun) 1 is absent and the crown wheel 2 is blocked (FIG. 7).
In this gear train, the crank OΩ forms the planet carrier 3 whereas the gearwheel with external toothing forms the pinion 4. From a kinematical viewpoint, the planet carrier 3 only rotates around its own axis (Oz), whereas the pinion or planet 4 is characterised by a composite motion, one motion consisting of a rotation around the axis through Ω, and the other, of a revolution around the axis passing through O, together with the planet carrier 3.
Considering two levorotatory reference frames Oxyz and Oξηz in which the first one is an absolute frame “integral” with the crown wheel 2 with internal toothing, and the second one is a relative frame “integral” with the planet carrier, their common axis z being perpendicular to the plane of motion, and imposing a rotation αt=αz to the planet carrier (and therefore to the reference frame Q with respect to the reference frame Oxyz, it follows that the planet 4—being obliged to mesh with a gearwheel with twice its pitch radius—will rotate by an angle αr=−2αz with respect to the planet carrier 3, that is, with respect to the relative reference frame Ωξηz; therefore, the angle of rotation of the planet 4 with respect to the absolute reference frame Oxyz will be αa=αr+αt=−2αz+αz=−αz. FIG. 8 shows various positions of the “non-conventional” crank mechanism of the background art, for various crank angles α. Assuming point B to be fixed to (“integral with”) the planet 4, the path (trajectory) of this point during the rotation of the planet carrier 3, in the absolute frame, will be a rectilinear segment. Point B can be embodied, in practice, by a pin and a bush, wherein the piston 5 may be connected to the planet 4 by a rod 6, attached to the piston without a hinge and on the planet 4 through said pin. As already mentioned, there are various known techniques which have put into practice the above described kinematical system; however, they were unsuccessful in practice since they offer technical solutions that have some inconsistencies and render impossible a correct operation, while in other cases they result in a great structural complexity which discourages their use.
The following list includes some filed patent applications based on the above operation principle:
U.S. Pat. No. 2,271,766 filed Feb. 3, 1942 of H. A. HUEBOTTER
U.S. Pat. No. 875,110 filed Apr. 30, 1953 of Harald Schultze, Bochum
U.S. Pat. No. 3,626,786 filed Dec. 14, 1971 of Haruo Kinoshita et al.
U.S. Pat. No. 3,791,227 filed Feb. 12, 1974 of Myron E. Cherry
Patent No. DE 36 04 254 A1 filed Feb. 11, 1986 of TRAN, Ton Dat
Patent No. DE 44 31 726 A1 filed Sep. 6, 1994 of Hans Gerhards
Italian Patent No. 1309063 of LAME S.r.l.
As a matter of fact, though they take advantage of a “non-conventional” crank mechanism which is undoubtedly better than the classical one (because of the above reasons), no one of the above mentioned patents has in practice been applied industrially to positive-displacement reciprocating compressors, notwithstanding the fact that some of these solutions of the background art seem to be valid; actually, often the structural complexity was excessive, space problems arose, and reliability was insufficient. These problems have rendered uncompetitive the “non-conventional” crank mechanism with respect to the classical one.
Therefore, it is desirable to provide a compressor that operates according to a “non-conventional” crank mechanism but which has—in contrast with the background art—the advantages of a reduced size, increased reliability, less components (resulting in a facilitated assembling and structural simplicity), together with self-lubrication features for the constituent material of the components in relative motion of the crank mechanism. Moreover, it is desirable to use a processing technique that lowers production costs of the mechanical parts of the “non-conventional” crank mechanism.
Furthermore, compressors of the background art obviously require that a certain amount of lubricant, usually oil, be fed to the components which are in relative motion.
To supply the necessary amount of liquid lubricant, compressors must be provided with lubrication systems capable of feeding even very modest lubricant flow rates but delivering them where they are actually needed; further, these lubrication systems must have a simple mechanics, low production costs, be capable of drawing the motion from the machine on which they are mounted without resorting to excessively complicated mechanisms (additional small shafts (spindles), power takeoffs, etc). At present, the lubrication of reciprocating compressors is essentially performed either by splash lubrication—provided this system reveals itself sufficient—or by means of gear pumps, if the needs of a good lubrication are more strict. Recently, electromagnetically controlled pumps, or small reciprocating, mechanically controlled pumps (generally based on cams) have also been devised, for instance in small-sized internal combustion engines for scooters or motorcycles. The present invention is a valid alternative to conventionally used solutions like those employed in the field of the lubrication systems for positive-displacement compressors.
The alternative proposed by the present invention consist of a lubrication system which, during operation, directly draws the mechanical energy necessary for its motion from the driving shaft of the compressor, and lubricates in an accurate (targeted) manner those components of the “non-conventional” crank mechanism—of the compressor—which are in relative motion with respect to each other. This system has a very convenient cost, it does not require power takeoffs or independent drive means, it is extremely easy to assemble, and it does not “waste” lubricant oil since it directs the latter exactly towards those parts which are in relative motion. It will be noted, in the detailed description of the invention, that in combination with a “non-conventional” crank mechanism provided with self-lubricant properties (due to its constituent material), this lubrication system, to be described later, will insure a perfect lubrication together with obvious economical advantages.
Although the classical splash lubrication of the background art, relying on the splashing and entrainment caused by the very components to be lubricated (which are wetted by the oil generally contained in an oil sump) has the advantage to be extremely economical and simple, provided it insures a sufficient lubrication, it has—nevertheless—considerable drawbacks, like the need to maintain a constant lubricant level inside the oil sump in order to avoid seizure. Moreover, in this way the lubricant is not accurately supplied (that is, it is not exclusively supplied to the points where it is really needed), since this system does not feed the lubricant under pressure. Moreover, this system cannot be employed in two-strokes engines with sump oil pump, since in these applications the sump oil pump must work under dry conditions.
Therefore, in the background art the lubrication under pressure has become the most widespread system because of its evident advantages linked to its utilisation, these advantages being, among others, the increase in performance of the kinematical couples lubricated under pressure as compared with that obtainable without the contribution of the feed pressure.
In particular, lubrication effected by gear pumps according to the background art offers the advantage of putting the lubrication circuit under pressure, thereby allowing to precisely reach the various points to be lubricated, with the correct oil flow rate and the required pressure. In that case the lubricant also has the not negligible task of cooling the surfaces which are in mutual contact. Also the use of cam-actuated reciprocating pumps has quickly become widespread, in the same way as electromagnetic pumps, in the field of small-sized internal combustion engines, due to the possibility of feeding the lubricant under pressure, by controlling the flow rates and therefore, taking advantage of the possibility of cooling down the various lubricated kinematical couples.
However, the disadvantage in the use of gear pumps lies in the increased cost involved in the production of high-quality mechanical components, like the gearwheels for instance, and in the need to provide an adequate power takeoff (drive), so that the machine to be lubricated will be more difficult to manufacture. On the other hand, the drawbacks of using cam-actuated pumps, in their commonly used version, are the requirement of their assembling in the vicinity of the driving shaft and the need of having available an adequate oil level in the oil sump in order to permit the priming (pump starting). The drawbacks of using electromagnetically controlled pumps are generally the increased production cost, their electric power absorption, and the necessity of providing a control unit.
Summing up, it would be desirable to provide a positive-displacement reciprocating compressor which has a targeted (accurate), economical, reliable, compact, and easily mountable lubrication system, which directly draws the power from the compressor drive shaft and does not require high-quality complex mechanical components that need complex machining for their production (case of gear pumps), and moreover, a lubrication system not requiring an excessive amount of oil in the oil sump.
A further problem of the background art relates to the system of intake valves (suction valves) and delivery (head) valves of a positive-displacement reciprocating compressor.
Compressor valves may be actuated mechanically or automatically; the first case covers for instance the valves that are actuated by means of cams; the second case includes the type of valves whose opening/closing is caused by the pressure difference existing between the upstream and downstream regions of the valve. Mechanical valves have the advantage of following a precise ‘lift law’ but their considerable disadvantage lies in the complex structure, the great number of auxiliary elements involved, the fact that they are excessively cumbersome, their weight and their cost. All these factors have determined a situation in which, practically, all commercial compressors used in conventional applications have been equipped with automatic valves. The commonly used automatic valve system is formed by (see FIGS. 9a, 9b, 10) two identical plates 7, 8 having appropriate seats to receive two flexible lamellar blades (which normally are made of harmonic steel). The plates 7, 8 are usually identical and are mounted face-to-face in an asymmetric manner, with the flexible lamellar blades located within these appositely realised seats, so as to form a single package, and so that the fluid flow directions allowed by the valves are opposite to each other. The valve package is usually mounted on a cylinder head of the compressor in such a manner that one side of this package directly faces the inner space of the cylinder, whereas the other side faces towards the cylinder head located above the cylinder. Normally,
the cylinder head is divided in two distinct regions isolated from each other by a sealed septum or dividing wall. A first of these regions is traversed by the suction or intake flow, while a second region is traversed by the delivery flow. The first region allows the flow to enter by virtue of the depression which, generated inside the cylinder as a consequence of the descending motion of the piston from the top dead centre to the bottom dead centre, causes the opening of the intake valve. The latter is shaped so as to allow the passage of working fluid from the outside of the cylinder to the inside of the same while preventing its passage in the inverse direction. The second region allows the fluid (which has been compressed in the cylinder by the piston during the ascending stroke from the bottom dead centre to the top dead centre) to exit from the cylinder after the opening of the discharge valve. The latter is shaped so as to allow the working fluid to pass from the inside to the outside of the cylinder, while blocking the inverse path. The opening of the valves therefore occurs as a consequence of the pressure difference on the two opposite sides (faces) of each lamellar blade. This pressure difference causes the inflection of the lamellar blades—which obviously behave in this case in the same way as simple beams supported at both ends and subjected to a distributed load—, thereby opening a passage for fluid flow which is directed from the upstream region to the downstream region with respect to the blades and their valve seats. These seats are in turn realised on the plates so as to allow the inflection (bending) of each lamellar blade to take place in one direction only, and for a limited, maximum opening (bending), in such a way that the blades immediately close when the pressure gradient that caused their opening changes sign. Thus, these valves, as has already been said, essentially act as check valves.
This automatic valve system of the background art is surely efficient, and with respect to that realised by means of mechanically actuated valves it is certainly more simple and economic; however, also this system has drawbacks. The first of them is due to the inevitable increase of clearances, consisting of volumes that correspond to the necessary passage areas obtained on the surface of one of these plates used to retain the lamellar blades, in particular of that plate which directly faces the inside of the cylinder, which adds to the volume of the seat (space) that receives the suction valve (see space 9 in FIG. 9b). The second drawback is the presence of two asymmetrically arranged plates 7, 8 which face each other and which contain the lamellar blades, and moreover, another drawback resides in the difficulty of assembling these components and in the often occurring overheating problems of the delivery lamellar blades, which are interposed between the plates and are therefore influenced by the high temperatures of the delivery flow, without being protected by an efficient thermal exchange that would limit the maximum temperature reached by them. In the FIGS. 9a and 9b there is shown, in an exploded view, the package (assembly) of plates according to this background art and according to a usual, commercially available embodiment, in the typical arrangement in which the cylinder (not shown) is located below the two plates. FIG. 9a corresponds to a lower-side view, while FIG. 9b is an upper-side view of the two plates 7, 8. Number 8 denotes the lower plate, number 7 the upper plate. The number 10 indicates the lower face of plate 8, which is the face facing towards the cylinder inside (not shown). On this lower face 10, in its middle part, there are rectangular slits. This group of four slits 11, located on the left, forms the slits traversed by the fluid which enters the compressor by passing beyond the suction valve, when the lamellar blade 12 that forms the latter (see FIG. 10) is open. The slit arranged on the right, denoted by the number 13 in FIG. 9a, is the one traversed by the outgoing flow of the compressor, when, during the compression stroke, the inner pressure overcomes the outside pressure and thereby determines the opening of the lamellar blade 14 (FIG. 10) which forms the delivery valve. FIG. 9b shows the holes 15 used for mounting the plates on the cylinder head. These holes 15 are formed on both of the plates 7, 8 to be connected together. In FIG. 9b one sees the upper view of this upper plate 7. On the left, the space or seat 9 is visible, which is occupied by the lamellar blade 12 made of harmonic steel forming the suction valve, while on the right one notes a slit 16 to be traversed by the fluid under pressure that exits the cylinder. FIG. 9a also shows the lower face of the lower plate 8. On the left, one notes a slit 17 to be traversed by the suction flow during the opening period of the corresponding lamellar blade 12, while on the right there is a space or seat 18 occupied by the lamellar blade 14 made of harmonic steel (FIG. 10), which forms the delivery valve. In FIG. 9b, finally, one notes the upper face of the upper plate 7, which shows a perfectly asymmetric arrangement with respect to the lower face of the lower plate 8 as already shown in the above mentioned FIG. 9a. One may note the slit 17, through which the sucked fluid passes when crossing the intake or suction valve, and holes 18, which are traversed by the compressed fluid when it leaves the delivery valve whose lamellar blade is denoted by 14 in FIG. 10.
An alternative system of automatic valves according to the background art is realised by resorting to a single plate having appropriate seats used to lodge the two flexible lamellar blades (one for the suction flow and the other for the discharge flow, usually of harmonic steel), both of these valves being—however—usually connected to the plate at one of their ends, so that their opening occurs only on one side, by a simple inflection. The connection is generally obtained by a rivet or another means suited to realise a stable connection with the plate.
Now, another object of the present invention, according to a more specific embodiment of the same included in the dependent claims, is obtained by means of a realisation which provides a particular valve system in the positive-displacement reciprocating compressor.
This object consists in providing a valve system in the positive-displacement reciprocating compressor, this valve system solving some of the problems which have been mentioned previously and which are inherent problems of known automatic valve systems (which are present both in single-stage compressors and multistage compressors).
In particular, the objects that can be attained by utilising a valve system according to the present invention, are the following—as will be detailed in the subsequent, more precise description of the invention—:
(case concerning a single-stage compressor or the first stage of a multistage compressor)
                A reduction of the clearance, since the noxious volume only concerns the delivery valve. In fact, the suction valve, being directly oriented towards the interior of the cylinder, does not “add any volume” to the clearance (on the contrary, it reduces it by a small amount);        A reduction in the number of components, since in this embodiment only one valve plate is necessary, unlike the traditional systems which employ two plates 7, 8. The component number reduction implies less machining work and reduced production costs;        A simplification in the assembling process because of the reduced number of components, and impossibility of an erroneous assembling/mounting;        A solution of the overheating problem for the delivery valve, because this valve, which is lodged in the delivery space or volume of the cylinder head, is no more forced to stay inside a very narrow volume surrounded by walls at high temperature. (case concerning a stage located downstream of the first stage in a multistage compressor)        A reduction of the clearance, since the noxious volume, or space, is the sum of that volume associated with the delivery valve (which is minimal, since the valve directly faces the cylinder) and of that volume which concerns the intake valve (which is minimal due to the fact that the valve seat (valve space) has been laterally arranged with respect to the outer edge of the cylinder);        A reduction of the number of components, since this embodiment only requires a single valve plate for the delivery valve, which also incorporates the components needed for the intake valve (suction valve), instead of the two plates of the traditional systems, which imply clearances that have a greater value, taken as a whole. Also in this case, the reduction of the number of components, and therefore the utilisation of a single plate, implies a reduction of the involved machining processes and related production costs;        A simplification of the assembling and mounting process, due to the utilisation of a single plate and a resulting impossibility of erroneous assembling;        A solution to the delivery valve overheating problem, since this valve, which is lodged inside the delivery volume of the cylinder head, is no more forced to stay inside a very narrow volume surrounded by walls at high temperature;        Elimination of the heat exchange (heat transmission) between the sucked fluid and the compressed fluid; this heat exchange taking place in traditional systems through the thin septum (dividing wall) located between the adjacent volumes present inside the cylinder head. In the proposed system, the sucked fluid is not subjected to such heating; which consequently reduces the work expended during compression.        